Geared transmissions

ABSTRACT

A gear assembly ( 10 ) for transmitting torque between an input rotatable member ( 12 ) and an output rotatable member e.g. differential unit ( 24 ) comprises an input gear ( 14 ) rotatable with the input member ( 12 ), an output gear ( 20 ) rotatable with the output member ( 24 ), and intermediate gears ( 16   a   , 16   b   , 16   c   , 16   d ) held simultaneously in mesh with the input gear ( 14 ) and output gear ( 20 ). Various arrangements are disclosed for sharing the transmitted torque evenly between the paths provided by gears ( 16   a   , 16   b   , 16   c   , 16   d ). The transmission may provide a reduction ratio and is connected to a low reduction ratio differential gearbox of small size and slim profile.

[0001] This invention relates to transfer gearing for transmitting powerbetween first and second members rotating about parallel axes inautomotive transmissions, for example in which power from the outputshaft of a main gearbox must be transmitted to a pair of propellershafts respectively extending fore and aft of the vehicle. Thesepropeller shafts are usually interconnected by differential gearing, andextend generally parallel to the main variable ratio gearbox outputshaft. Typically, power is transmitted between a gear on the maingearbox output shaft and a similarly sized differential gear, to give anapproximate 1:1 gear ratio. There are other instances where power mustbe transmitted from a first shaft to a second, generally parallel shaftin an automotive transmission.

[0002] In heavy vehicles such as trucks, this power transmission isoften via an intermediate gear supported on a layshaft of a transfergearbox. The diameter and/or lateral offset of the intermediate gear maybe selected to provide the required centre spacing between the parallelshafts. Where a large centre spacing is required, the intermediate geardiameter must be made correspondingly large. As the maximumtransmissible torque depends mainly upon the load per unit tooth width,the intermediate gear and the co-operating gears must each be maderelatively wide for transmission of high torque loads.

[0003] High torque loads and large centre spacings thus lead to a largeand heavy gear train between the main and transfer gearboxes. Largediameter gears have a higher pitch line velocity and hence tend to benoisier in operation than smaller gears.

[0004] For lighter 4×4 “sports utility” vehicles, such transfer boxesvirtually all use chain drives for power transmission between theparallel shafts. The driveline must operate at high RPM and because ofthe necessary gear diameter/operating centre distances a conventionalsingle power path geared drive as used in trucks, for example, wouldinvolve unacceptably large gears and pitch line velocities. At thespeeds concerned, chain drive has proven quieter than even precisionground gears and allows a more compact casing. However chain drivessuffer from overheating problems and excessive centrifugal loadings ifused at speeds above about 6000 RPM. Their power transmission capacityis therefore limited.

[0005] Theoretically, an alternative design approach would be to providea pair of intermediate gears offset to either side of the planecontaining the parallel shaft axes, these intermediate gears eachmeshing simultaneously with corresponding gears on the parallel shafts(e.g. in the main and differential gearboxes respectively). Assumingthat all four such gears are perfectly concentric, with perfect toothpitches and profiles, supported on shafts perfectly spaced relative toone another, journalled in perfect, play-free bearings, the whole beingmade from perfectly inelastic materials, the transmitted torque will beshared equally between the two intermediate gears, to provide paralleltorque transmission paths. The intermediate gears and the co-operatinggears in the main and differential gearboxes could thus theoretically bemade correspondingly smaller and lighter.

[0006] However, commercially manufactured gearboxes are not perfect. Thekinematic forces acting on the intermediate gears coupled with theelasticity of the materials of the gear assemblies means that in realityone of the intermediate gears tends to be forced inwards towards theplane of the input/output shafts, whilst the other intermediate geartends to be forced outwards away from that plane. The intermediate gearforced inwards experiences a higher torque than the intermediate gearforced outwards.

[0007] Furthermore, dimensional inaccuracies in the various gearboxcomponents means that in reality one of the intermediate gears will,when torque is applied, assume flank-to-flank drive contact with each ofthe two adjacent gears, whilst at that instant the opposite intermediategear has not established drive contact. Thus at that time only onetorque transmission path is effective. As the torque load increases,provided that dimensional inaccuracies are within acceptable limits,gearbox components will deform under load until mutual drive contact isestablished between all adjacent gears. However, torque sharing betweenthe two transmission paths will be unequal, with the degree ofinequality corresponding to the size of the dimensional inaccuracies.The proportion of the torque transmitted through each path may varythroughout the rotation cycle of the gearbox assembly, as thedimensional inaccuracies of each gear may vary cyclically.

[0008] Studies by NASA on helicopter gearboxes (see paper by Timothy L.Krantz “A Method to Analyze and Optimize the Load Sharing of Split-PathTransmissions”, published in Design Engineering, vol. 88, PowerTransmission and Gearing Conference ASME 1996, at pages 227-242) haveshown that satisfactory torque sharing between two parallel transmissionpaths can be achieved if tooth flank position errors are controlled toless than 0.0005 radian. Under these conditions, the inequality oftorque transmission might not vary beyond, say, 60:40, leading toworthwhile savings in gearbox size and weight. Such dimensional accuracyis achievable, certainly in aerospace and similar specialistapplications where high manufacturing costs are not prohibitive. Howeverthe requirement for high dimensional accuracy means that such torquesharing arrangements are impractical for mass produced automotivegearboxes, where low cost is an important factor.

[0009] U.S. Pat. No. 6,035,956 discloses an axle for low platform townbuses in which hub reduction gear trains are connected, one on eachside, between the axle differential and respective offset stub axlescarrying the bus road wheels. Each transfer gear train comprises a pairof intermediate gears providing parallel power transmission paths. Aninput gear fixed to a respective output shaft of the differential mesheswith both intermediate gears simultaneously and is vertically movable soas to share torque evenly between the power transmission paths.

[0010] U.S. Pat. No. 5,896,775 concerns high reduction final drivegearing for a powered scooter or wheelchair, in which an input shaft isconnected to drive a pair of ground wheels through a pair of torquesharing pinion gears. The pinion gears engage a further gear wheelconnected to drive the ground wheels. In one embodiment, the furthergear wheel contains a differential arrangement.

[0011] It has now been realised that plural power path arrangementsincorporating even torque sharing capability are of significant benefitto transfer gearing elsewhere in automotive transmissions, in particularbetween the main variable ratio gearbox and the axle (differential)drive, and also in other locations “up stream” of the axle differential.

[0012] In accordance with the invention there is provided an automotivetransmission comprising a transfer gear train for transmitting torquebetween an input rotatable member and an output shaft rotating aboutsubstantially parallel axes, the transfer gear train comprising an inputgear rotatable with the input member, an output gear rotatable with theoutput shaft, and a pair of intermediate gears each held simultaneouslyin mesh with the input gear and transmitting torque to the output gearto provide two power transmission paths, characterised in that theoutput shaft drives differential gearing arranged to distribute drivingtorque to a pair of ground engaging wheels. Preferably, one of the gearsin the transfer train is made movable in response to the transmittedtorque so as to even out power transmission between the two paths.However such torque sharing can also be achieved by other means, such asby controlling gear tooth flank position errors to within acceptably lowlimits.

[0013] The input gear is preferably made smaller than the output gear sothat the transfer gear train provides a reduction ratio. Thedifferential gearing may therefore have a lower reduction ratio, evensubstantially 1:1. This enables it to be made considerably smaller andlighter. Pitch line velocities in the transfer gearing and in the restof the transmission driven by it are also reduced, giving quieteroperation. The sizes of the gears in the transfer gear train anddifferential can be smaller than in a conventional single power pathdriveline, which much reduces gear weight (roughly proportional todiameter squared) and very much reduces gear moments of inertia (roughlyproportional to diameter cubed).

[0014] In current driveline designs for passenger cars, the axle bevelgear has a large diameter, typically providing a reduction ratio ofabout 3:1—even larger for heavy vehicles. Any diminution in thisreduction ratio increases gear loadings “up stream”, including in themain gearbox and transfer gearbox (if present). Current drivelineproportions therefore represent a trade off between minimum gearbox sizeand maximum acceptable axle ratio. The plural power path transfer geartrain of the present invention provides a high capacity, compact powertransfer path that can handle the higher torque loads arising from theuse of low ratio axle differentials, and may itself be used to provide areduction ratio, thereby reducing torque loadings on-the main gearbox.

[0015] A smaller axle differential gives a greater ground clearance,which is important in off-road vehicles. Incorporating speed reductionin the transfer gear train also enables the overall reduction ratio ofthe transmission to be maintained, whilst using a smaller reductionratio at the differential gearbox. This is beneficial in highperformance vehicles such as racing cars, with engines operating at highRPM and which therefore require a high overall driveline reductionratio. For such applications, the transfer gear train of the presentinvention again gives lower pitch line velocities and smaller, lighter,quieter, lower inertia differential gearboxes and final drivelines.

[0016] The two power paths enable the transfer gear train to be madesmaller and lighter than a conventional transfer gearbox of equivalentduty. To facilitate assembly of the various gears in proper meshingengagement and to provide flexibility in the available gear ratios andin the centre spacing between the input and output members, a furtherintermediate gear is preferably provided in each power transmissionpath.

[0017] The torque responsive movement referred to above may for examplebe of the input gear. In one possible arrangement, the input shaft andthe rotational axes of the co-operating intermediate gears all liesubstantially in a common plane.

[0018] Where the various gears are spur gears, the input shaft may bemade free to move (e.g. pivot) very slightly away from this plane. Then,when there is flank to flank contact between the teeth of the input gearand only one of the intermediate gears, the torque applied to the inputshaft and the reaction at the contacting tooth flanks will form a couplecausing the input shaft to move out of the common plane. This movementcontinues until there is flank to flank contact at the otherintermediate gear as well. Even torque sharing between the two powertransmission paths is therefore achieved.

[0019] Where the various gears are single helical gears, the input gearmay be free to pivot about an axis normal to the common plane.Out-of-balance forces acting in the direction of the input shaft axisand arising from uneven torque sharing between the two power paths willrotate the input gear about the pivot axis in a direction tending toreduce the out-of-balance forces, and hence evening up the torquesharing. Operation of such a mechanism is more fully explained in GB1434928. To reduce or substantially eliminate thrust loads on the gearshafts, the gears may be mounted to their shafts by helical splinedconnections, the helix being of the same hand and having the same leadas the gear teeth. The splined connection may itself allow the inputgear to pivot about an axis passing through the input gear.

[0020] Similar arrangements are possible in which even torque sharing orcompensation is provided by movement of the output gear.

[0021] The torque compensating gear may also be free to move slightlyboth in a direction normal to the common plane and/or along the axis ofthe input rotatable member, as described for example in relation to theoutput gears 24, 34 in EP 0244263. Another arrangement permitting suchtranslational movements of a torque compensating gear is describedbelow.

[0022] The invention in its various aspects, and its further preferredfeatures, are described below with reference to illustrative embodimentsshown in the drawings, wherein:

[0023]FIG. 1 is a diagrammatic front view of a transfer gear trainembodying the invention;

[0024]FIG. 2 is a diagrammatic sectional view of an automotivetransmission incorporating the transfer gear train of FIG. 1;

[0025]FIG. 3 is a diagram showing a mounting arrangement for a torquecompensating single helical gear that may be used in embodiments of theinvention;

[0026]FIG. 4 is a diagrammatic section of a further differential unitwhich may be used in a modification of the FIG. 2 transmission for fourwheel drive vehicles, and

[0027]FIG. 5 is a diagrammatic section of a planetary differential unitwhich may be used in a further modification of the FIG. 2 transmission.

[0028] Referring to FIG. 1, a transfer gear train 10 for an automotivetransmission, comprises an input rotatable member in the form of aninput shaft 12 connected to the output of a variable ratio main gearbox(not shown). Input shaft 12 carries an input gear 14 which meshessimultaneously with a pair of first intermediate gears 16 a, 16 bmounted on intermediate shafts 18 a, 18 b journalled in a gearbox casing(not shown). The intermediate gears 16 a, 16 b mesh with respectivefurther intermediate gears 16 c, 16 d mounted on shafts 18 c, 18 d. Thefurther intermediate gears 16 b, 16 c mesh simultaneously with an outputgear 20. In the drawings, the circles illustrate the pitch lines of thevarious gears shown. For simplicity, the gear teeth are not illustrated.

[0029] As shown in FIG. 2, the output gear 20 is fixed to the inputshaft 26 of an axle differential unit 34, having output shafts 36 a, 36b for driving respective road wheels (not shown). The differential unit34 may be of any known conventional kind, for example incorporating sliplimiting or lockup means.

[0030] To provide even torque sharing between the parallel power pathsconstituted by the gears 16 a, 16 c on the one hand and 16 b, 16 d onthe other, the input shaft 12 is mounted so that its end carrying thegear 14 is movable slightly out of the plane containing the shafts 18 a,18 b. The bearings 30 may allow pivoting of the shaft 12 about an axisnormal to the page (as indicated by the arrows 32) and the shaft 12 mayincorporate a splined or other connection for this purpose. Such torquecompensating movement is suitable for use with spur gears. If insteadthe gears 14, 16 a, 16 b, 16 c, 16 d are single helical, gear 14 may bemounted to the shaft 12 in the manner described in GB 1434928 to providethe necessary torque compensating movement. Similarly, if the variousgears are double helical, mounted in “herringbone” configuration, torquecompensation of the resulting compound gear 14 can be as disclosed in EP0244263.

[0031]FIG. 3 shows a further torque sharing arrangement for use withsingle helical gears. The compensating movement is provided by mountinggear 14 to shaft 12 via a torque transmitting sleeve 40. Gear 14 isformed as a ring gear having internal helical splines engagingcomplementary splines 42 on the sleeve 40. Internal helical splines onthe sleeve in turn engage complementary splines 44 on the shaft 12. Thesplines 42, 44 are short and slightly crowned, to allow pivoting of theaxes of shaft 12, sleeve 40 and gear 14 relative to each other. The leadand hand of the splines 42, 44 are equal to the lead and hand of thehelical gear teeth 46. This ensures that axial forces on the sleeve 40and gear 14, arising from the transmitted torque, balance out. With thisarrangement, the gear 14 is not only free to pivot slightly out of theplane normal to the shaft 12, but can also translate slightly, bothaxially and normal to the plane passing through the axes of shafts 18 aand 18 b. This provides improved torque sharing with respect to themechanism of GB 1434928.

[0032] Although in the drawings the torque sharing or compensationmechanism is shown applied to the input gear 14, it could equally beapplied to an output or other gear meshing simultaneously with a pair offurther gears, the shafts of all three gears being substantiallyco-planar. For example, the torque compensating movement may be of theoutput gear, particularly in transfer gear trains providing a step-upratio.

[0033] As shown, the input gear 14 is smaller than the output gear 20.This therefore provides a reduction ratio. The reduction ratio of thefinal drive differential gearbox 34 may therefore be made smaller. Asmaller bevel gear 38 may therefore be used. This results in a muchslimmer gearbox 34, further reducing the weights and amounts ofmaterials required, and improving the vehicle ground clearance. Largeoverall transmission speed reduction ratios may also be achieved, whichcan be advantageous in high performance vehicles such as Formula 1 GrandPrix racing cars. These currently use servo-operated, clutchlessgearshifts, the gearbox having close ratios so as to eliminate therequirement for synchronisers. A large reduction ratio is needed in thetransmission final drive. The main gearbox output shaft is close to theground, so that the drive must be taken upwards to the rear wheels. Thearrangement shown in FIG. 2 is suitable for such use. (For clarity, FIG.2 shows the main transmission elements in co-planar configuration. Whenused in racing car transmissions, the gear train 10 may extend upwardly,i.e. the transfer gear train 10 is rotated from the position shown,relative to the axle differential 34, about the shaft 26, so as to placethe shaft 12 at a lower level than the shaft 26).

[0034] A large centre spacing between the shafts 12, 26 is possible,whilst keeping the size of the gears 16 a, 16 b, 16 c, 16 d reasonablysmall. The transfer gearbox is accordingly compact, lightweight, quiet,efficient, has low inertia, and is capable of handling high shaftpowers. Still further intermediate gears can be added into the powertransmission paths, as desired.

[0035] It is also possible to eliminate the further intermediate gears18 c, 18 d to produce a four gear arrangement in which the gears 18 a,18 b mesh directly with gear 20 (the centres of gears 18 a, 14 and 18 bstill being substantially in line to provide torque sharing). However,the gear ratios must then be carefully selected so as to achieve propermeshing of the gear teeth and so as to avoid clashing of gears 14 and20. Only a limited number of gear ratios and offsets between shafts 12and 26 are therefore available, which can be found e.g. by numericalmeans. Some examples are tabulated below. Numbers of teeth Gears GearsGears Gear 18a, Gear Gear 18a, Gear Gear 18a, Gear 14 18b 20 14 18b 2014 18b 20 5 13 23 15 27 57 15 267 297 5 19 29 15 33 63 30 72 132 5 25 3515 45 75 30 108 168 5 41 31 15 51 81 30 120 180 5 37 47 15 63 93 30 144204 5 43 53 15 69 99 30 156 216 5 49 59 15 81 111 30 180 240 5 55 65 1587 117 30 192 252 5 61 71 15 99 129 30 216 276 5 67 77 15 105 135 30 228288 5 73 83 15 117 147 30 252 312 5 79 89 15 123 153 30 264 324 5 85 9515 135 165 30 288 348 10 20 40 15 141 171 30 300 360 10 32 52 15 153 18330 324 384 10 44 64 15 159 189 30 336 390 10 56 76 15 171 201 30 360 41410 68 88 15 177 207 30 372 432 10 80 100 15 189 219 30 396 456 10 92 11215 195 225 30 408 468 10 104 124 15 207 237 30 432 492 10 116 136 15 213243 30 444 504 10 128 148 15 225 255 30 468 528 10 140 160 15 231 261 30480 540 10 152 172 15 243 273 30 504 564 10 164 184 15 249 279 30 516576 10 176 196 15 261 291 30 84 144

[0036]FIG. 4 shows a further differential unit 24 for use in four wheeldrive (or higher) transmissions. Instead of being fixed directly to theshaft 26, the gear 20 is formed as a ring gear attached to adifferential carrier 50 rotatable in a casing 52 of the differentialunit 24. The shaft 26 forms one of the output shafts of the differentialunit 24. A further output shaft 28 takes the drive to a further pair ofvehicle road wheels (not shown), via a further axle differential unit(not shown), similar to unit 34. The transmission arrangement isotherwise similar to FIGS. 1 and 2, with the ring gear 20 engagedsimultaneously by the gears 16 c, 16 d (only gear 16 c is visible inFIG. 5; in “four gear” transfer train arrangements, ring gear 20 is ofcourse engaged directly by gears 16 a, 16 b). The differential unit 24splits the input torque equally between the output shafts 26, 28. Likethe axle differential 34, it may incorporate conventional slip limitingand lock-up means. The transfer gear train 10 may be rotated about theshafts 26, 28, from the position shown, to any desired configuration.

[0037]FIG. 5 shows a modification of the FIG. 4 arrangement, using aplanetary differential 60. The ring gear 20 has external teeth forengagement by the gears 16 c, 16 d (or 16 a, 16 b) as in FIGS. 1, 2 and4. It also has internal teeth engaged by a plurality of (e.g. three)planet gears 62, which in turn engage a sun gear 64. The gears 62 arejournalled in a planet carrier 66, which drives one output shaft 26. Thesun gear 64 drives the other output shaft 28. The planetary differential60 provides an uneven torque split between the shafts 26 28 in the ratio

[0038] d_(s)+d_(p)/2: d_(s),

[0039] where d_(s) is the diameter of the sun gear 64 and d_(p) is thediameter of the planet gear 62. The differential 60 may be provided withotherwise conventional lock-up or slip limiting means, acting betweenthe planet carrier 52 and sun gear 64.

1. An automotive transmission (10, 26, 28, 34) comprising a transfergear train (10) for transmitting torque between an input rotatablemember (12) and an output shaft (26, 28) rotating about substantiallyparallel axes, the transfer gear train (10) comprising an input gear(14) rotatable with the input member (12), an output gear (20, 62, 64)rotatable with the output shaft (26, 28), and a pair of intermediategears (16 a, 16 b) each held simultaneously in mesh with the input gear(14) and transmitting torque to the output gear (20, 62, 64) to providetwo power transmission paths; characterised in that the output shaft(26, 28) drives differential gearing (34) arranged to distribute drivingtorque to a pair of ground engaging wheels.
 2. A transmission as definedin claim 1 characterised in that the input gear (14) is smaller than theoutput gear (20).
 3. A transmission as defined in claim 1 or 2characterised in that one of the gears in the transfer train is mademovable in response to the transmitted torque so as to even out powertransmission between the two paths.
 4. A transmission as defined inclaim 3 characterised in that the torque responsive movement is of theinput gear (14).
 5. A transmission as defined in any preceding claim,characterised in that a further intermediate gear (16 c, 16 d) isprovided in each power transmission path.
 6. A transmission as definedin any preceding claim, characterised in that the rotational axis one ofthe gears (14) in the gear train which is movable in response to thetransmitted torque so as to even out power transmission between the twopaths, and the rotational axes of two other gears (16 a, 16 b) in thegear train which mesh simultaneously 30 with the movable gear (14), alllie substantially in a common plane.
 7. A transmission as defined inclaim 6 characterised in that it comprises spur gears (14, 16 a, 16 b,16 c, 16 d, 20) and wherein the shaft (12) of the movable gear (14) isfree to move away from the common plane.
 8. A transmission as defined inclaim 6 characterised in that it comprises single helical gears (14, 16a, 16 b, 16 c, 16 d, 20) and the movable gear (14) is free to pivotabout an axis normal to the common plane.
 9. A transmission as definedin claim 8 or 9 characterised in that the movable gear (14) is mountedto a shaft (12) by a helically splined connection (42, 44), the helix ofthe splines being of the same hand and having the same lead as the helixof the gear teeth.
 10. A transmission as defined in any of claims 6-9characterised in that the movable gear is free to translate out of thecommon plane and/or along its rotational axis.
 11. A transmission asdefined in claim 10 characterised in that the movable gear (14) ismounted to the shaft (12) via a torque transmitting sleeve (40), therebeing helically splined connections (44, 42) between the shaft (12) andthe sleeve (40) and between the sleeve (40) and the movable gear (14).12. A transmission as defined in any of claims 1 to 5, characterised inthat one of said gears (14) comprises two single helical gears mountedto a common shaft by helical splined connections, the common shaft beingcoupled to the input (14) or output (26) shaft by a drive connectionthat allows axial and radial displacement of the common shaft.